EP1338884B1 - Roller test bench for motor vehicles - Google Patents

Description

The invention relates to a chassis dynamometer for motor vehicles with two rotatable roller strands, the uprights have at their respective axial ends, the wheels of an axle of the vehicle respectively on a pair of rollers consisting of a foot roll of the first roller strand and a foot roll of the second roller strand stand up (Doppelrollenprüfstand).

To test driving dynamics parameters and the emission values of motor vehicles, chassis dynamometers are used, which enable a simulation of different driving conditions. At least the driven wheels of the vehicle stand on the crest area of each a riot roll or in the role prism of a pair of rollers. The uprising rolls or the pairs of rolls are coupled to drive / deceleration machines, which correspond to the Test program generate predetermined torques or tensile forces to measure operating parameters of the vehicle, such as the performance of the internal combustion engine, or to check the functions of components, such as brakes or emission control system. During the simulation on the test bench, predetermined driving cycles are run through and acceleration or deceleration of the vehicle is simulated by the introduced torques or forces.

To adapt the rotating rotating masses of the test rig to different vehicle masses different flywheel sets, with which the vehicle masses are simulated, are coupled by couplings to the roller strands. The actual rotational masses of the rollers are usually relatively small compared to the dominant flywheels. The vehicle mass or its mass moment of inertia should correspond to the translational test bench mass that corresponds to the rotating mass of the test bench or its mass moment of inertia. For an exact adaptation of the test bench to the vehicle mass, the difference between the vehicle mass and the mechanical mass gradations of the flywheel sets can additionally be compensated by an electrical mass simulation.

wobei die Beschleunigung der Rolle aus $\frac{\mathrm{\Delta }v}{\mathrm{\Delta }t}$

bestimmt wird.In the DE 199 00 620 a mass simulation method is described in which an acceleration to be simulated force F for the difference mass m between the mechanical Prüfdredrehmasse (mass moment of inertia) and the vehicle mass from the second Newton's law (F = ma) is derived: $$\mathrm{F}\mathrm{=}\mathrm{m}\cdot \frac{\mathrm{\Delta }v}{\mathrm{\Delta }t}$$

being the acceleration of the role $\frac{\mathrm{\Delta }v}{\mathrm{\Delta }t}$

is determined.

Since the masses to be simulated by the coupling of the flywheels are usually relatively small (less than 100 kg), any torsional vibrations occurring during the rotation of the roller strands, which impair a speed measurement for the rollers, do not lead to any major problems in the measurement of the roller acceleration ,

mAfter US 5 101 660 The second Newtonian law can also be used to determine the simulated driving speed: $$\mathrm{v}\mathrm{=}\frac{\mathrm{F}\mathrm{\cdot }\mathrm{t}}{\mathrm{m}}\mathrm{,}$$

m

In this method, the driving speed is calculated from the integral of the force taking into account the mass m to be electrically simulated. Possibly occurring torsional vibrations in the rotation of the roller strands have no disturbing influence during integration.

In order to avoid the mechanically complex coupling of different flywheel sets, only a single mechanical flywheel is used in newer test stands. The remaining mass differences are covered by an electrical mass simulation to adapt the rotating test bed mass to the vehicle mass. In such a test stand, for example, the mechanical translational mass of the test stand is 1,500 kg. This is adapted by the electrical mass simulation of vehicle masses, for example, from 300 to 3,000 kg. Compared to mechanical mass matching, it is now necessary to electrically simulate masses of, for example, up to 1,500 kg. This places correspondingly high demands on the sensors of the test stand, in particular on the detection of the speed or the acceleration. Possibly occurring torsional vibrations can affect the accuracy of the speed measurement and - amplified by the differentiation of the speed to calculate the acceleration - affect the force to be simulated. Since the force to be simulated or the corresponding torque is an important variable in the control of the roller drives, control problems can thus occur due to the torsional vibrations.

It is particularly advantageous and simple in construction, if you provide the complete basic inertia of the test in the roles and renounces additional flywheels.

In dual roller test stands, the wheels run in the roller prism of a pair of rollers and the fixation of the vehicle is less complicated than in the case of crown roller test stands. Since the Aufstandsrollen in a Doppelrollenprüfstand usually have a significantly smaller roll diameter (eg 500 mm), while common apex rollers have a large diameter (eg 1200 mm) and a larger flywheel, the electrical mass simulation for Doppelrollenprüfstände is more expensive and makes increased demands on the test sensors. For example, the four individual rollers of a double roller dynamometer each have a translatory mass of about 350 kg (JX = 21.9 kg / m ² with a roller diameter of 500 mm). The total translational mass of the test rig is then 1,400 kg. The differential mass to vehicle mass is electrically simulated by applying Newton's second law. Now occur torsional vibrations, ie, the masses of the rollers oscillate against each other, an accurate mass simulation is compromised due to the inaccuracies in the speed measurement.

After the conventional construction of a double roller dynamometer in the so-called inline variant of the drive motor is mounted in extension of a first roller string with two uprights for the left and right lane of the vehicle. In the FIG. 6 a conventional structure for a double roller dynamometer is shown. The drive motor drives the first roller train directly and the second roller train via a mechanical roller strand connection, usually a belt or a chain. The shafts of the first and second roller string are usually rotatably supported by a plurality of bearings. Due to the resulting bearing friction losses, which are also not constant and depend, for example, on air pressure and bearing temperature, it is difficult to determine the exact forces or torques for accurate simulation of driving resistance, since the temporally changing bearing friction counteracts the engine torque at non-compensated bearings. Uncompensated bearings are outside the measuring or control chain. This leads to unstable Prüfstseigenverlusten that affect the accuracy of the test bench.

In the conventional construction of a double roller dynamometer, strong torsion or torsional vibrations can also occur due to a lower torsional rigidity of the shafts. Also, the mechanical pulley connection between the first and second roller strand, the spring-mass system of the roller strands is affected by the elasticity of the belt or the chain and the formation of torsional vibrations is favored. Due to the torsional vibrations, the speed measurements are disturbed, which can lead to problems in the control of the rotational speed and / or the acceleration of the rollers. Measurement inaccuracies in the detection of the roller speed are also amplified when differentiating to determine the roller acceleration. As for the simulation of driving resistance forces acting on the vehicle's traction is proportional to the roller acceleration, the accuracy of the driving simulation or the mass simulation is impaired by the torsional vibrations.

Since an electrical simulation of large difference masses is required in particular in Doppelrollenprüfständen with relatively low translational mass of the rollers, torsional vibrations between the masses of the roller strands have a disturbing influence. In the lateral coupling of the drive unit, for example, vibrations occur between the two masses (rollers) of the front roller string against each other, wherein the shaft of the roller strand is the elastic element. By the coupling of the two roller strands also creates an elasticity between the front and rear rotating masses (for example, each 700 kg translational) of the roller strands, which causes a vibration of the two roller strands against each other. Such a chassis dynamometer is from the document EP-A-0 424 636 known.

Another disadvantage of the inline engine assembly is the large lateral space requirement. In particular, for mobile test stands, which are mounted for example on a trailer or truck, which are not much wider than the motor vehicle to be tested, it is difficult to arrange the drive motor in extension of a roller string.

The object of the present invention is to provide a compact double roller dynamometer with reduced friction losses, in which small torsional vibrations occur and an accurate simulation of driving resistance forces or of rotating masses is possible.

This object is achieved by the independent claim. The dependent claims relate to advantageous embodiments of the invention.

A motor vehicle dynamometer according to the invention has two rotatable roller strands with contact rollers and a roller train shaft connecting the contact rollers. The contact rollers are preferably arranged on the respective axial end regions of the roller strands. The wheels of an axle of the vehicle are in each case on a pair of rollers consisting of a contact roller of the first roller strand and a contact roller of the second roller strand. The two roller strands are preferably arranged in parallel so that the vehicle wheels run in the roller prism of the contact rollers. The uprising rollers of a roller string can be fixedly connected to each other by a rotatable shaft or coupled to the shaft via one or more couplings. The diameter of the tread rolls is preferably 200 to 600 mm, with common rolls having a diameter of 20 inches or 500 mm. In the Abgasmeßtechnik roll diameters of 220 to 550 mm are common. The distance between the two rollover roles of a roll strand to each other and the width of the riot rolls are sized so that vehicles with different track widths can stand up on the uprising roles. For testing all-wheel drive vehicles, two such roller sets may be provided for the two axles of the vehicle.

According to the invention, at least one drive unit is provided for driving or decelerating at least one roller train and a control unit for the drive unit. The drive unit can drive or decelerate the land rolls to induce forces to simulate driving resistances on the wheels of the vehicle. During the driving resistance simulation, test conditions predetermined by the vehicle on the chassis dynamometer are passed through to determine, for example, the exhaust emissions of the vehicle or the like. The drive unit is suitably designed as a DC motor, three-phase asynchronous motor or three-phase synchronous motor. In particular, synchronous machines can be advantageously used to generate a predetermined torque or a predetermined speed using frequency converters.

In order to simulate selected driving states, in a preferred embodiment of the invention, the control device, taking into account specific vehicle data, controls the drive unit in order to release corresponding loads in the form of tractive forces on the vehicle wheels. The corresponding tractive forces can be used to simulate different driving resistances, for example for gradients or gradients. The control unit may further comprise input and output units for operation and for data communication and be realized for example by a process computer or a programmable controller.

According to the invention, the drive unit is arranged in a region which is bounded by the axially inner ends of the contact rollers. Due to the internal arrangement of the drive unit, a very compact construction of the chassis dynamometer is possible since no laterally projecting beyond the support rollers drive or connecting devices are provided. The inventive arrangement of the drive unit of the dynamometer can be mounted, for example, on a trailer or truck, which are not much wider than the vehicles to be tested.

Since the internal arrangement of the drive unit initiates the driving force on the roller strands between the contact rollers, a torsion-resistant structure is created in comparison with the conventional in-line arrangement. Advantageously, the introduction of the drive torque takes place centrally in the roller train. The maximal Path length from the location of the force on the roller strand to the location of the load of the roller strand is inventively reduced, resulting in a reduced spring effect of the shaft and to improvements in the properties of the spring-mass system of the roller strand. If, according to the invention, the driving force is coupled centrally into the roller string, the two rotating masses of the roller strand do not oscillate in opposite phase. The inventive arrangement of the drive unit, the rotational or torsional vibrations of the rotating masses of the roller strands are reduced, which among other things leads to more stable speed measurements, an increase in the measurement accuracy for the acceleration measurement and an improvement of the control dynamics. In particular, the regulation of the acceleration for the electrical mass simulation is thereby improved.

An advantageous embodiment of the invention provides to couple the drive unit via a centrally between the spacer rollers on the roller strand arranged belt connection, in particular a toothed belt with the roller strand. The drive torque can be introduced in this way centrally in the roller train to reduce torsional vibrations of the contact rolls. The drive unit may be arranged in the area bounded by the contact rollers in the roll strand longitudinal direction. It is advantageous if the drive unit for reasons of space is not arranged directly between the support rollers, but is located in the test stand longitudinal direction in front of or behind the roller strands.

Due to the compact design of the chassis dynamometer according to the invention, it is possible, in the axial direction immediately next to the support rollers other devices, such as holding devices for the vehicle to install. Due to the internal arrangement of the drive unit of the dynamometer according to the invention with less uncompensated bearings and thus be realized with reduced or compensatable Prüfstseigenverlusten.

A further advantageous embodiment of the invention provides to arrange the drive unit between the two contact rollers of the first roller strand to initiate the driving force directly to the shaft of the first roller strand. Furthermore, it is advantageous if the axis of rotation of the contact rollers of the driven roller train coincides with the axis of rotation of a rotor shaft of the drive unit, since in the case of a common axis of rotation particularly favorable vibration and torsional properties for the roller strand and this is particularly simple. A particularly advantageous embodiment of the invention provides that the contact rollers are attached to the axial ends of a one-piece, continuous rotor shaft of the drive motor. When using an electric motor as a drive unit, the rotor shaft between the contact rollers on the rotor windings of the electric motor. A one-piece design of the roller string shaft is very stable and stiff, so that little torsion and deflection occurs. By a symmetrical introduction of the (pendulum) moments in the mass system of the roller strand occur no torsional vibrations between the masses (roles) of the roller string. A compensation of the torsional vibrations is not required.

A roller dynamometer according to the invention preferably has at least one measuring device for determining the rotational speed and / or the angular position of the roller strand. Based on the detected measurement signal, the control unit can regulate the rotational speed and / or the angular position of the insurrection wheels. To measure the rotational speed and / or the angular position, sensors can be attached to the contact rollers and / or the roller string shafts. Swing into the roller string, especially with centered force input the two rotating masses of the roller strand are not in opposite phase, whereby it is possible to detect the speed or acceleration of the roller train with a sensor at one end of the shaft. The measurement can be carried out by means of incremental encoders or so-called sine / cosine encoders.

In order to compensate vibration frequencies up to 30 Hz in the correct phase, it is expedient to carry out the determination of the acceleration with a speed which is higher by a factor of 10, since the acceleration is used for mass simulation and for regulating the drive units. The determination of the acceleration of the riser rollers is expediently two hundred to three hundred times per second by differentiating the detected speed. In order to achieve the necessary accuracy of the calculated acceleration, it is advantageous if the speed is detected with an accuracy of at least 14 bits. In order to achieve the required high temporal resolution and accuracy in speed detection, it is advantageous to use sine / cosine encoders.

To improve the speed measurement, a plurality of measuring devices may be provided on any chassis dynamometer for motor vehicles whose measuring signals are superimposed, in particular by addition or averaging. By an opposite phase superposition of interfering signals they can be suppressed and it is possible to more accurate speed measurement. The measuring devices can be mounted at different locations of a roller train or on different roller strands such that interference signals occur in anti-phase on the measurement signals. The measuring signals can also be appropriately phase-shifted added to suppress interference. By a corresponding arrangement, in particular the low-frequency modulation of the speed signal (about 10 Hz) be reduced by the torsional vibrations in the roller strands. As a result of the improved rotational speed or angular position signals, the roller acceleration can be determined more accurately, which allows a more precise simulation of the driving resistance forces or the differences.

An inventive dynamometer for motor vehicles has at least one rotatable roller train with pad rollers, at least one drive unit for driving or decelerating the roller train, a control unit for the drive unit and at least two measuring devices for determining a rotational speed and / or an angular position of the roller strands, wherein at least the wheels of a Stand up axle of the vehicle on the support rollers and the control unit for suppressing disturbances in the determined rotational speed and / or angular position takes the signals of the measuring devices into account.

In a method for suppressing disturbances in rotational speed and / or angular position signals of motor vehicle chassis dynamometers, a plurality of rotational speed and / or angular position signals are detected for the at least one rotatable roller strand, on whose uprising wheels the wheels of the vehicle stand up, and optionally added out of phase.

Conveniently, at least one force or torque measuring device is provided which determines the force applied by the drive unit or the torque. For the force or torque measurement, for example, the power consumed by the drive unit can be used. Preferably, a mechanical force or torque sensor, such as a load cell or a strain gauge, may be provided.

In a preferred embodiment of the invention, a mechanical pulley connection, in particular a belt, a toothed belt or a chain, is provided for driving the second roller strand. By the roller string connection, the rotational movement of the second roller string is mechanically coupled to the rotational movement of the driven first roller string to achieve a synchronous rotation of the pad rollers. It may be particularly important to accurately synchronize the rotational movements of the respective uprights of the pairs of rollers in order to prevent slippage between the vehicle and the corresponding Aufstandsrollenpaar. To achieve favorable torsion properties, a plurality of roll strand connections can be provided. Preferably, a symmetrical division of the power flow from the driven roller strand to the second roller strand by the provision of two roller string connections, which are each arranged at the axially inner end of a roller pair. The arrangement of the roll strand connections can be asymmetric or symmetrical to the test bench central axis. A symmetrical arrangement of the roller strand connections, for example symmetrical belt connections between the roller strands, leads to the reduction of torsional vibrations between the rotating masses of the rollers. To compensate for the torsional vibrations of the first roller strand to the second roller strand, a high-resolution speed sensor for the roller strands can be used in each case, wherein the detected speeds or the calculated accelerations of the roller strands are averaged arithmetically in phase to compensate for the influence of torsional vibrations.

A further advantageous embodiment of the invention provides to connect the two roller strands via a common chain or a common belt, in particular a toothed belt, with the drive unit. The joint connection directs the drive torque preferably in the middle of the roller strands in this one to reduce torsional vibrations. The drive assembly is conveniently located in the region between the roller strands and below the roller strands to provide a triangular arrangement for the belt connection with the roller strands. It may also be desirable to provide a double belt connection between the roller strands and the power plant to achieve uniform power transmission and reduced torsional vibration. Preferably, two belt connections are provided symmetrically to the test bench longitudinal axis, wherein the drive unit is arranged between the two belt connections and drives both belt connections. The shaft of the drive unit may each have a pulley at both ends, on which rests the respective (toothed) belt.

To further reduce torsional vibrations, the flywheel mass of the idler rollers of the driven roller string may be larger than the flywheel of the idler rollers of the second roller train. Due to the uneven mass distribution, the spring-mass system of the test bench and thus the vibration characteristics of the system is affected. Furthermore, the spring constant of the roll strand connection, in particular the coefficient of expansion of a belt, and the arrangement of the roller strand connection (symmetric / asymmetric) have an effect on the vibration characteristics of the test bed system. To improve the vibration characteristics of the test bench and to reduce the interference on the speed measurement, for example, 70% of the masses of the rollers in the driven roller train and 30% in the second roller train, can be provided. By deliberately influencing the vibration characteristic, it is also possible to influence interference by irregularities or vibrations and their transmission be reduced between the roller strands by the coupling of the roller strands.

In a particularly preferred embodiment of the invention, two drive units are provided which each drive a roller train. Each drive unit, for example, an electric three-phase machine, accelerates or decelerates the associated role strand. It is advantageous if the respective uprising rollers of the roller strands are mounted directly on the ends of the continuous rotor shafts of the drive units. In this two-engine concept, a mechanical roller strand connection can be omitted. The spring-mass system of this embodiment has no appreciable elastic elements. Due to the lack of mechanical coupling of the roller strands, the torsional vibrations of the test stand are reduced. Furthermore, the Prüfstseigenverluste, especially under changing environmental conditions, and the control dynamics are improved and increases the stability of synchronization. Through the use of two drive units, the performance of the test stand can be further increased or it is possible to provide drive units with lower power. This is particularly advantageous because there are limitations in the size of the drive units in the center motor design of the test stand according to the invention.

Each roller string may have a measuring device for determining the rotational speed and / or the angular position of the roller strand, in order to enable independent detection of the movement variables of the roller strands. The control unit advantageously controls the rotational speeds and / or the angular positions of the two roller strands such that both roller strands have the same rotational speed and / or angular position. Since the synchronous synchronization of the contact rolls by an electrical synchronization control Without mechanical pulley connection takes place, a very high stability of the Prüfstseigenverluste is achieved.

In particular, for the examination of four-wheel drive vehicles, the roller dynamometer can also have two sets of two roller strands, wherein all four driven vehicle wheels stand up in each case on a pair of rollers. The required for the examination of four-wheel vehicles synchronization of all insurrection wheels can be done by an electronic synchronization control of the drive units.

If an exact synchronization synchronization of the roller strands is not required, provides a cost-effective variant of the invention that the control unit has a common actuator that drives both drive units. In this case, both motors are operated with a common power unit, that is, for example, that two parallel-connected electric motors are supplied with the same control signals. In this case, the speed control can adjust a predetermined value for the rotational speed of a roller strand or take into account the mean value of the rotational speeds of the two roller strands as a controlled variable.

Preferably, the shaft of a roller strand is rotatably supported by means of suitable rolling bearings. In order to determine the forces acting on the vehicle wheels as accurately as possible, low and constant as possible bearing friction losses are required. It is therefore advantageous to heat the bearings in order to achieve constant test standstill losses even under changing environmental conditions, as e.g. occur in a cold chamber.

A particularly advantageous embodiment of the invention provides to store the stator of the drive unit oscillating rotatable. For the mutual storage of the stator housing in the test rig frame the stator housing aurweisen on each end face a hollow pin, in which via inner roller bearings, the rotor shaft is mounted. Opposite the test stand frame each hollow pin is mounted on a further radially outer roller bearing. The torque exerted by the drive unit on the contact rollers can be detected as a reaction torque of the stator housing by a torque or force sensor. The friction of the rotor shaft bearing is thus detected by the measuring device and can be automatically compensated. The rotor shaft bearings are located within the measuring or control chain. Thus, even when changing the operating modes (continuous operation, short-time operation) or different environmental conditions (eg temperature, air pressure) the Prüfstseigenverluste be detected and compensated.

To detect the Statorreaktionsmoments the stator is preferably supported via a torque arm and a Krafmeßeinrichtung on the test rig frame. From the detected force and the dimensions of the torque arm can be determined in a simple manner the torque exerted by the pendulum rotatably mounted drive unit on the roller strand. As a force measuring device, a load cell, which is preferably supported floating on the frame of the test stand, are used. Due to the floating support of the load cell measuring accuracy is increased because friction and tilting can be avoided. The measured torque or the determined force can be used to control the drive unit and / or to calibrate / adjust the control unit and / or to compensate for the rotor shaft bearing friction. Furthermore, various vehicle properties, such as braking force or drive power, based on the detected torque and the force can be determined.

To avoid wear of the bearing parts in the oscillating mounting of the stator, the stator can be mounted with counter-driven bearings. Here, two roller bearings are arranged coaxially with the interposition of a middle ring,
wherein the center ring of the arranged on the two end faces of the drive unit bearing rotate in opposite directions to compensate for the moment generated by the rotation of the center rings. The drive for the rotating center rings can either be done individually or have a single drive machine for all center rings, which drives the center rings, for example via a belt drive.

In order to detect the bearing friction of the second, non-driven roller strand, it is expedient to pivotally support the shaft of the second roller strand. In this pendulum rotatable mounting can, similar to the oscillating mounting of the stator housing, the outer bearing shells of the rotor shaft bearings are in a housing which is mounted on an outer bearing in the test rig frame. To avoid bearing damage in the outer, stationary bearing this can also be performed as a double bearing with rotating center ring. By a mounted on the bearing housing torque sensor, the torque of the second roller strand can be detected. Since both the bearing friction of the first driven roller train and that of the second roller train can be detected, no undefined bearing friction is present in the test bench and it is possible to compensate for the friction losses of the total test bench. This can be done by a corresponding detection of the friction loss for the respective test conditions and an adjustment / calibration of the control unit or by controlling the desired torques by the control unit.

A further embodiment of the invention provides that the pendulum rotatably mounted stator of the second drive unit or the pendulum rotatably mounted shaft of the second roller train acts on the force measuring device via a second torque arm. Due to the mechanical connection of the second stator housing or of the bearing housing of the second roller train with the force measuring device, a mechanical addition of the forces to the load cell can take place. By a suitable design of the torque arms bearing friction losses of the entire test bench can be detected and automatically compensated.

In an advantageous embodiment of the chassis dynamometer, the bearings of the roller strands are arranged between the support rollers. Due to the arrangement of the drive unit in a region between the contact rollers, it is not necessary to provide bearings at the axially outer ends of the roller strands. Thus, the number of bearings and the associated bearing friction losses can be reduced. It is also possible to realize a particularly compact test bench with small dimensions.

The inventive arrangement of the drive unit, a compact chassis dynamometer with reduced friction losses and less tendency to torsional vibrations can be achieved. Furthermore, it is possible to achieve a trouble-free detection of the roller speeds or speeds and thus a more accurate calculation of the roller accelerations for the simulation of driving resistances and / or of rotating masses. The roller dynamometer according to the invention also makes it possible to detect and compensate for the variable bearing friction losses.

Fig. 1

eine schematische Darstellung eines Doppelrollenprüfstands in Mittelmotorausführung;

Fig. 2

eine schematische Darstellung eines Doppelrollenprüfstands mit pendelnd drehbarer Lagerung des zweiten Rollenstrangs;

Fig. 3

eine schematische Darstellung eines Doppelrollenprüfstands mit zwei Antriebsaggregaten;

Fig. 4

eine schematische Darstellung eines Schnitts durch einen Rollenstrang mit pendelnd drehbarer Lagerung des Statorgehäuses;

Fig. 5

eine schematische Darstellung eines Schnitts durch einen Rollenstrang mit doppelter Lagerung des Statorgehäuses; und

Fig. 6

eine schematische Darstellung eines herkömmlichen Doppel-rollenprüfstands.

Other features and advantages of the invention can be taken from the drawings and the following description of preferred embodiments. Show it:

Fig. 1

a schematic representation of a double roller dynamometer in mid-engine design;

Fig. 2

a schematic representation of a double roller dynamometer with pendulum rotatable mounting of the second roller strand;

Fig. 3

a schematic representation of a double roller dynamometer with two drive units;

Fig. 4

a schematic representation of a section through a roller train with pendulum rotatable mounting of the stator housing;

Fig. 5

a schematic representation of a section through a roller train with double storage of the stator housing; and

Fig. 6

a schematic representation of a conventional double roller dynamometer.

The Fig. 1 shows a schematic representation of a double roller dynamometer in mid-engine design with a drive unit. The chassis dynamometer has a first roller string 1 with contact rollers 3, 4 and a second roller string 2 with contact rollers 5, 6. The wheels of an axle of the vehicle to be tested are each on a pair of rollers 3, 5 or 4, 6. After driving the vehicle to the test stand, the vehicle wheels are in the respective roller prism of the roller pairs and are accelerated or decelerated by the rotating rollers 3, 4, 5, 6. To fix the vehicle on the test bench are usually in Fig. 1 Not shown holding devices provided. For testing all-wheel drive vehicles, it is possible to provide two roller sets each with two roller strands.

wobei folgendes gilt:

F:

Zugkraftsollwert

f₀:

konstanter Koeffizient (Rollwiderstandsanteil)

f₁:

linearer Koeffizient (Walkwiderstandsanteil)

f₂:

exponentieller Koeffizient (Luftwiderstandsanteil mit n = 2)

f₃:

exponentieller Koeffizient (Exponent variabel)

n:

Exponent (1 < n < 3, eine Nachkommastelle)

m_g:

Statische Fahrzeugmasse (Fahrzeuggewicht)

v:

Rollengeschwindigkeit

dv/dt:

Rollenbeschleunigung

g:

Erdbeschleunigung

sinα:

Steigungswinkel (+).

The simulation of driving resistance forces on the vehicle wheels takes place according to the following equation, $$\mathrm{F}\mathrm{=}{\mathrm{f}}_{\mathrm{0}}\mathrm{+}\left({\mathrm{<figure-callout\; id="1"\; label="f"\; filenames="imgf0001.png"\; state="\{\{state\}\}">f</figure-callout>}}_{\mathrm{1}}\mathrm{*}\mathrm{v}\right)\mathrm{+}\left({\mathrm{<figure-callout\; id="2"\; label="f"\; filenames="imgf0001.png"\; state="\{\{state\}\}">f</figure-callout>}}_{\mathrm{2}}\mathrm{*}{\mathrm{<figure-callout\; id="2"\; label="v"\; filenames="imgf0001.png"\; state="\{\{state\}\}">v</figure-callout>}}^{\mathrm{2}}\right)\mathrm{+}\left({\mathrm{<figure-callout\; id="3"\; label="f"\; filenames="imgf0001.png,imgf0002.png"\; state="\{\{state\}\}">f</figure-callout>}}_{\mathrm{3}}\mathrm{*}{\mathrm{v}}^{\mathrm{n}}\right)\mathrm{+}{\mathrm{m}}_{\mathrm{G}}\mathrm{\cdot }\left(\mathrm{dv}\mathrm{/}\mathrm{dt}\right)\mathrm{+}\left({\mathrm{m}}_{\mathrm{G}}\mathrm{*}\mathrm{G}\mathrm{*}\mathrm{sin\; .alpha}\right)$$

where:

F:

Traction setpoint

f ₀ :

constant coefficient (rolling resistance component)

f ₁ :

linear coefficient (full resistivity)

f ₂ :

exponential coefficient (aerodynamic drag with n = 2)

f ₃ :

exponential coefficient (exponent variable)

n:

Exponent (1 < n < 3, one decimal place)

m _g :

Static vehicle mass (vehicle weight)

v:

roller speed

dv / dt:

role acceleration

G:

acceleration of gravity

sin .alpha:

Pitch angle ( + ).

The roll diameters of the tread rolls 3, 4, 5, 6 are preferably 20 inches (508 mm), 500 mm or 220 to 550 mm in the exhaust gas measuring technique. To measure the exhaust emissions of the vehicle this is subjected to a predetermined test cycle with defined driving resistance forces on the test bench, wherein the development of exhaust gas of the vehicle is detected by an exhaust gas measuring device.

To drive or delay the contact rollers 3, 4, 5, 6, a drive unit 7 is provided, which is arranged between the two contact rollers 3, 4 of the first roller strand 1. Due to the central force or torque introduction into the roll strand 1 Torsional vibrations between the contact rollers 3, 4 reduced. The uprising rollers 3, 4 are attached to the axial ends of a rotor shaft 17 of the electric motor 7. The motor 7 for driving the first roller strand 1 is preferably an electric three-phase motor, since this allows easy control of the speed and the torque. For torque measurement, a torque measuring device 9 is attached to the pendulum rotatably mounted drive unit 7. By measuring the torque acting on the rotor shaft 17 from the motor 7, a particularly accurate simulation of driving resistance forces can take place. Furthermore, acceleration or deceleration forces, which are introduced by the vehicle wheels on the contact rolls, can be detected and the test stand can also be used for brake testing or power measurement. Of course, other uses for the dynamometer according to the invention are possible.

A speed measuring device 8 is provided on a contact roll 3 in order to determine the rotational speed and / or the angular position of the roller strand 1. Based on this speed signal, a control unit 16, the rotational speed and / or the angular position of the roller strand 1 regulate.

By differentiating the speed, the roller acceleration for the electrical mass simulation can also be determined. On the basis of the second Newton's law, a force is determined, which is applied by the drive unit 7 via the contact rollers 3, 4, 5, 6 on the vehicle wheels to simulate a difference between the mechanical test inertia and the vehicle mass. Since the rotating masses of the insurrection rolls or their moments of inertia in relation to the other rotating masses of the test bench dominate, the latter can in the electrical mass simulation be largely neglected. Due to the large masses to be simulated in a Doppelrollenprüfstand with relatively small uprights high demands on the accuracy of the speed or Geschwindigkeitsmeßeinrichtung 8. In order to achieve an accurate and temporally high-resolution speed measurement, therefore preferably sine / cosine encoders are used. The calculation of the acceleration takes place at least two hundred times per second by differentiating the detected speed.

The shaft 26 of the second shaft train 2 is rotatably supported by roller bearings 11, while the storage of the first roller train 1 provides a pendulum rotatable mounting 10 of the motor stator to detect the friction losses of the engine bearings by the measuring device 9 and compensate automatically. By compensating the test bench inherent losses, precise simulations of driving resistance forces under changing operating conditions are possible.

Mechanical roller assemblies 12, 12a connect the first roller string 1 to the second roller assembly 2. The roller assembly 12, 12a can be made, for example, by a belt or a chain and drive the second roller assembly 2 in synchronism with the rotational movement of the first roller assembly 1. The symmetrical connection between the roller strands torsional vibrations between the roller strands 1, 2 are reduced.

The Fig. 2 shows a schematic representation of a chassis dynamometer with pendulum rotatable mounting of the shaft 26 of the second roller strand 2. The bearing of the shaft 26 of the second roller strand 2 can be done similar to the oscillating mounting of the stator of the motor 7. A housing 13 receives the inner bearings 18 of the roller string shaft 26 and is over the outer bearing 10 a stored floating. To avoid bearing damage this outer bearing 10 a, this may be formed by a double bearing with a rotating center ring between the two bearings. To compensate for the forces and moments occurring through the rotating center rings of the double coaxial rolling bearings, it is advantageous if the middle rings are driven in opposite directions. The drive of the rotating center rings can be done for example via belts that transmit a rotational movement of an auxiliary motor. The oscillating mounting of the stator of the drive unit 7 can be done by means of counter-driven bearing.

For detecting the total torque of the two roller strands 1, 2 or the entire test bench losses a torque arm 14 is provided, which acts on the measuring device 9. In addition to this mechanical addition of the forces or torques, it is of course also possible to provide a second measuring device for detecting the torque or the bearing friction of the second roller strand 2 and to carry out an electronic calculation of the detected moments or forces.

By providing a second roller strand connection 12, which is provided symmetrically to the test bench central axis, a symmetrical distribution of the power flow and thus a reduction of the torsional vibrations can be achieved. The vibration characteristic of the test stand can be further improved by an asymmetric mass distribution, in order to additionally reduce the torsional vibrations. A favorable asymmetric mass disposition provides, for example, to arrange 70% of the masses of inertia of the insurgency rolls in the first roll strand 1 and 30% of the momentum masses in the second roll strand 2. The asymmetric mass distribution reduces torsional vibrations and reduces the influence of vibrations and imbalances or irregularities on the speed measurement.

In the in Fig. 2 shown test stand, a second speed sensor 8a is further provided, which is arranged such that any interference occurring in phase opposition to the two speed signals occur. By superimposing the two speed signals, the speed measurement can be improved and stabilized. In particular, by attaching the second rotational speed sensor 8a at the other axial end of the first roller strand 1, the influence of any remaining torsional vibrations between the rotating masses of the contact rollers 3, 4 in the first roller strand 1 is reduced.

To compensate for the torsional vibrations of the first roller strand 1 to the second roller train 2, a third speed sensor 8b is provided on the contact roller 5 of the second roller strand 2. The sensor 8b detects the rotational speed of the second roller train 2, which is arithmetically averaged in the control unit 16 with the rotational speed of the first roller train 1 detected by the first rotational speed sensor 8. For the electrical simulation of masses, it is also possible to first calculate the acceleration of the two roller strands 1, 2 and then to average them arithmetically.

The Fig. 3 shows a schematic representation of a double roller dynamometer with two drive units 7, 15, each driving a roller train 1, 2. Each roller strand 1, 2 has its own measuring device 8, 8b for determining the rotational speed and / or the angular position of the roller strand. Based on the signals of the measuring devices 8, 8b controls the control unit 16, the rotational speeds and / or angular positions of the two roller strands 1, 2 such that both roller strands 1, 2 have a same rotational speed and / or angular position.

To compensate for any oscillations between the rotating masses (contact rollers) of the two roller strands 1, 2, the detected speeds or the calculated accelerations of the roller strands can be arithmetically averaged. The mean roller acceleration can then be used for an accurate mass simulation.

The stators of the two drive units 7, 15 are each rotatably mounted by means of bearings 10 pendulum. A torque arm 14 connects the stator of the second motor with the torque measuring device 9. The mechanical torque addition, the total torque of the drive units 7, 15 are determined. Alternatively, separate torque measuring devices for the two roller strands 1, 2 may be provided. In this case, it is also possible to control the torques or forces of the individual roller strands 1, 2 separately.

Since in this embodiment, no mechanical roller strand connection is required, and the synchronization of the support rollers 3, 4, 5, 6 realized by an electronic synchronization control in the control unit 16, both roller strands 1, 2 are mechanically decoupled, whereby the system has a lower tendency to vibration. In particular, lower torsional vibrations occur because an elastic coupling between the roller strands 1, 2 is missing, the driving forces are introduced centrally into the roller strands and the maximum distance between the introduction of force into the roller string shafts 17, 26 and the load of the roller string shafts 17, 26 is reduced. Through the use of the second drive unit 15, the entire performance of the test bench can also be increased.

The Fig. 4 schematically shows a section through a roll strand. The contact rollers 3, 4 are attached to the axial ends of the continuous, one-piece rotor shaft 17 of the drive unit 7.

The oscillating rotatable mounting of the stator housing 19 is effected by the outer roller bearings 20 which support the stator housing 19 in the test bed frame 23. The rotor shaft 17 is rotatably supported in the stator housing 19 by the inner roller bearings 18. The rotor windings 21 of the electric drive unit 7 are fixedly connected to a roller string shaft 17. The stator windings 22 are located on the inside of the stator housing 19. The stator housing 19 is supported on the test stand frame 23 via a torque arm 24 and a load cell 25. The reaction torque of the stator housing 19 is detected by the force measuring device 25. The rotor shaft is integrally formed with the roller string shaft 17, which leads to a torsion and low bending bearing both the drive unit 7 and the back up rollers 3, 4.

The Fig. 5 shows a schematic representation of a section through a roller train with double, driven storage. In this, a further rolling bearing 28 is arranged coaxially over the bearing 20 to avoid bearing damage from stationary bearings and a center ring 27 is driven between these bearings 20, 28. In this way it can be avoided that the rolling elements dig into the bearing shells or tearing off the lubricant film in the rolling bearings.

The Fig. 6 schematically shows a conventional double roller dynamometer with an in-line arrangement of the drive unit 7 outside the support rollers 3, 4, 5, 6. Due to the unfavorable vibration characteristics of the spring-mass system such a test bench tends to increased torsional vibrations, which in turn lead to unstable speed measurements. Furthermore, a variety of bearings 11, which result in large and unstable test standings losses. Due to the lateral arrangement of the drive unit 7, a conventional test bed has a much greater width than the vehicle to be tested and therefore can not be mounted on a truck or trailer.

The described measures can also be used together with other chassis dynamometers to reduce torsional vibrations or to improve the speed signals.

Claims (21)

1. A roller dynamometer for motor vehicles, including
   two rotatable roller strings (1, 2) with footing rollers (3, 4, 5, 6), wherein the wheels of an axle of the vehicle each rest on a roller pair consisting of a footing roller (3, 4) of the first roller string (1) and a footing roller (5, 6) of the second roller string (2),
   at least one drive assembly (7, 15) for driving or decelerating at least one roller string (1, 2), and
   a controller (16) for the drive assembly (7, 15), wherein the drive assembly (7, 15) is disposed in a region bordered by the axially inboard ends of the footing rollers (3, 4, 5, 6), wherein the drive assembly (7) is connected to and drives or decelerates at least one roller string (1, 2) via an in particular centrally disposed belt connection.
2. The roller dynamometer according to claim 1, wherein the drive assembly (7) is disposed between the two footing rollers (3, 4) of the first roller string (1).
3. The roller dynamometer according to claim 3, wherein the rotation axis of the footing rollers (3, 4) of the driven roller string (1) coincides with the rotation axis of a rotor shaft of the drive assembly (7).
4. The roller dynamometer according to any one of claims 1 to 4, wherein at least one roller string has a measuring device (8) for determining the rotational speed and/or the angular position of the roller string and the controller (16) regulates the rotational speed and/or the angular position.
5. The roller dynamometer according to claim 5, wherein a plurality of measuring devices (8, 8a, 8b) are provided, the signals of which are added up or averaged.
6. The roller dynamometer according to any one of claims 1 to 6, wherein at least one force or torque measuring device (9) is provided, which determines the force or the torque, which is applied by the drive assembly (7).
7. The roller dynamometer according to any one of claims 1 to 7, wherein at least one roller string connection (12, 12a), in particular a belt or a chain, is provided for driving the second roller string (2).
8. The roller dynamometer according to claim 8, wherein the gyrating mass of the footing rollers (3, 4) of the driven roller string (1) is greater than the gyrating mass of the footing rollers (5, 6) of the second roller string (2).
9. The roller dynamometer according to any one of claims 1 to 7, wherein two drive assemblies (7, 15) are provided, which each drive a roller string.
10. The roller dynamometer according to claim 10, wherein each roller string (1, 2) has a measuring device (8, 8a, 8b) for determining the rotational speed and/or the angular position of the roller string (1,2).
11. The roller dynamometer according to claim 11, wherein the controller (16) regulates the rotational speeds and/or the angular positions of the two roller strings (1, 2) such that both roller strings (1, 2) have an equal rotational speed and/or angular position.
12. The roller dynamometer according to claim 10, wherein the controller (16) has a common actuator driving both drive assemblies (7,15).
13. The roller dynamometer according to at least one of claims 1 to 13, wherein the shaft (26) of a roller string is rotatably supported by means of rolling bearings (11).
14. The roller dynamometer according to at least one of claims 1 to 13, wherein a stator (19) of the drive assembly (7, 15) is oscillatingly rotatably supported.
15. The roller dynamometer according to claim 15, wherein the stator (19) is supported by means of driven bearings (10), in particular oppositely driven bearings.
16. The roller dynamometer according to claim 15 or 16, wherein the stator (19) is supported through a torque support (24) and a force measuring device (25).
17. The roller dynamometer according to at least one of claims 1 to 17, wherein the shaft (26) of the second roller string (2) is oscillatingly rotatably supported, in particular with oppositely driven bearings.
18. The roller dynamometer according to claim 17 or 18, wherein the oscillatingly rotatably supported stator (19) of the second drive assembly (15) or the oscillatingly rotatably supported shaft (26) of the second roller string acts on the force measuring device (9) through a second torque support (14).
19. The roller dynamometer according to at least one of claims 1 to 19, wherein the bearings (10, 11) of the roller strings (1, 2) are disposed between the footing rollers (3, 4, 5, 6).
20. The roller dynamometer according to at least one of claims 1 to 20, wherein a one-piece roller string shaft (17) is provided, which constitutes the rotation axis of the footing rollers (3, 4) and coincides with the rotor shaft.
21. The roller dynamometer according to at least one of claims 1 to 21, wherein a mass simulation for adapting the rotating dynamometer mass to the mass of the motor vehicle to be tested is effected by accelerating the roller strings (1, 2).

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